Multiple function hydraulic system with a variable displacement pump and a hydrostatic pump-motor

ABSTRACT

A hydraulic system includes a first pump and a plurality of valves that control fluid flow from the first pump to several actuators. Variable source orifices in the control valves are connected in parallel between the first pump and a node, and variable bypass orifices in the control valves are connected in series between the node and a tank. Pressure at the node controls displacement of the first pump. Each control valve also has a metering orifice for varying fluid flow between the node and one of the actuators. A hydrostatic pump-motor, coupled between two ports of a given actuator, is driven in a motoring mode by fluid exiting one of those ports. In a pumping mode, the hydrostatic pump-motor forces lower pressure fluid exhausting from one port into the other port of the given actuator.

CROSS-REFERENCE TO RELATED APPLICATIONS

This application claims benefit of U.S. provisional patent application no. 61/452,885 filed on Mar. 15, 2011.

STATEMENT REGARDING FEDERALLY SPONSORED RESEARCH OR DEVELOPMENT

Not Applicable

BACKGROUND OF THE INVENTION

1. Field of the Invention

The present invention relates to hydraulic systems having a plurality of pumps and a plurality of independently controllable hydraulic actuators; and more particularly to controlling the plurality of pumps and allocating the resultant fluid flow to the plurality of hydraulic actuators.

2. Description of the Related Art

Hydraulic systems have at least one hydraulic pump that supplies pressurized fluid which is feed through control valves to drive several different hydraulic actuators. A hydraulic actuator is a device, such as a cylinder-piston arrangement or a hydraulic motor that converts the flow of hydraulic fluid into mechanical motion.

Because loads of different magnitudes act on the various hydraulic actuators, the hydraulic pressure required to operate each actuator can vary greatly at any point in time. On an earth excavator, for example, the hydraulic actuators that raise the boom typically require a relatively high pressure as compared to other actuators that curl the bucket or move the arm. Thus, when the operator is raising the boom at the same time the arm or bucket are also moving, a significant portion of the fluid flow from the pump will go to the lower pressure hydraulic actuators. Without some further compensation mechanism, this deprives the boom actuator of the necessary fluid required to operate as commanded. To maintain the proper flow sharing among all the actuators, the hydraulic systems use complex throttling mechanisms that add a pressure drop to the lower pressure functions and prevent them from consuming a disproportionately large amount of the fluid flow at times when multiple actuators are operating. Different equipment manufacturers use different throttling mechanisms. Some of these mechanisms use pressure compensators and a load sensing pump, while other ones use pilot pressure signals from the operator controls to create throttling losses for the low pressure functions. All these throttling losses generate heat and add inefficiency to the hydraulic system in order to enable the multifunction operation commanded by a machine operator.

It is desirable to avoid these intrinsic losses in efficiency and energy while still maintaining the multifunction performance desired by the operator.

SUMMARY OF THE INVENTION

A hydraulic system for a machine includes a variable displacement first pump that furnishes pressurized fluid into a supply conduit in order to operate a plurality of hydraulic actuators. A return conduit is provided to receive fluid flowing back to a tank from the plurality of hydraulic actuators.

A control valve assembly has a plurality of control valves, each associated with a different one of the plurality of hydraulic actuators. Each control valve comprises a metering orifice for varying fluid flow between the first pump and the associated hydraulic actuator.

The hydraulic system further comprises a hydrostatic pump-motor connected to convey fluid between two ports of a given hydraulic actuator. In a motoring mode, the hydrostatic pump-motor is driven by fluid flowing out of one of the actuator ports at a higher pressure than fluid flowing into the other one of the ports. The hydrostatic pump-motor operates in a pumping mode when fluid flowing out of one port has a lower pressure than is required for fluid enter the other port.

One embodiment of the present hydraulic system further comprises a flow summation node that is in fluid communication with a control port for controlling displacement of the first pump. The metering orifice in each control valve varies fluid flow between the summation node and the associated hydraulic actuator. Each control valve further comprises a variable flow source orifice, and a variable bypass orifice. For example, each control valve is preferably configured so that as the metering orifice enlarges, the variable flow source orifice also enlarges and the variable bypass orifice shrinks. Inversely in that configuration, as the metering orifice shrinks, the variable flow source orifice also shrinks and the variable bypass orifice enlarges.

In the control valve assembly, the variable flow source orifices of the plurality of control valves are connected in parallel between the first pump and the flow summation node. The variable bypass orifices of plurality of control valves are connected in series to form a bypass passage through which fluid flows between the flow summation node and the return conduit.

BRIEF DESCRIPTION OF THE DRAWINGS

FIG. 1 is a pictorial view of an excavator having a hydraulic system;

FIG. 2 is a diagram of a first hydraulic system for the excavator;

FIG. 3 is a schematic functional diagram of the components of the first hydraulic system in FIG. 2 that control the displacement of a primary pump;

FIG. 4 is a diagram of a second hydraulic system for the excavator; and

FIG. 5 is a diagram of a third hydraulic system for the excavator.

DETAILED DESCRIPTION OF THE INVENTION

The term “directly connected” as used herein means that the associated components are connected together by a conduit without any intervening element, such as a valve, an orifice or other device, which restricts or controls the flow of fluid beyond the inherent restriction of any conduit. If a component is described as being “directly connected” between two points or elements, that component is directly connected to each such point or element.

Although the present invention is being described in the context of use on an earth excavator, it can be implemented on other types of hydraulically operated equipment.

With initial reference to FIG. 1, an excavator 10 comprises a cab 11 that can swing clockwise and counter-clockwise on a crawler 16. A boom assembly 12, attached to the cab, is subdivided into a boom 13, an arm 14, and a bucket 15 pivotally attached to each other. A pair of hydraulic piston-cylinder assemblies 17, that are mechanically and hydraulically connected in parallel, raise and lower the boom 13 with respect to the cab 11. On a typical excavator, the cylinder of these assemblies 17 is attached to the cab 11 while the piston rod is attached to the boom 13, thus the force of gravity acting on the boom tends to retract the piston rod into the cylinder. Nevertheless, the connection of the piston-cylinder assemblies could be such that gravity tends to extend the piston rod from the cylinder. The arm 14, supported at the remote end of the boom 13, can pivot forward and backward in response to operation of another hydraulic piston-cylinder assembly 18. The bucket 15 pivots at the tip of the arm when driven by yet another hydraulic piston-cylinder assembly 19. The bucket 15 can be replaced with other work heads.

With additional reference to FIG. 2, a pair of left and right bidirectional travel motors 20 and 22 independently drive the tracks 24 to propel the excavator over the ground. A bidirectional hydraulic swing motor 26 selectively rotates the cab 11 clockwise and counterclockwise with respect to the crawler 16.

The hydraulic motors 20, 22 and 26 and the hydraulic piston-cylinder assemblies 17-19 on the boom assembly 12 are generically referred to as hydraulic actuators, which are devices that convert hydraulic fluid flow into mechanical motion. A given hydraulic system may include other types of hydraulic actuators.

With particular reference to FIG. 2, a hydraulic system 30 has six hydraulic functions 31-36 although a greater or lesser number of such functions may be used in other hydraulic systems that practice the present invention. Specifically there are left and right travel functions 31 and 32 and a swing function 33. The boom assembly includes a boom function 34, an arm function 35, and a bucket function 36, referred to as implement functions.

Each hydraulic function 31, 32, 33, 34, 35 and 36 respectively comprises a control valve 41, 42, 43, 44, 45 and 46 and the associated hydraulic actuator 20, 22, 26, 17, 18 and 19. The six control valves 41-46 combine to form a control valve assembly 40. The control valves may be physically separate or combined in a single monolithic assembly. Control valves 41-46 govern the flow of fluid between the associated hydraulic actuator and both a variable-displacement primary pump 50 and a tank 51. The primary pump 50 furnishes pressurized fluid to a primary supply conduit 58 and the fluid return to the tank 51 through a return conduit 60. The primary supply conduit 58 and return conduit 60 or 110 extend to each of the control valves 41-46.

The primary pump 50 is of a type such that the output pressure is equal to a pressure applied to a load sense port 39 plus a fixed predefined amount referred to as the “pump margin”. The primary pump 50 increases or decreases its displacement in order to maintain the desired pressure. As an example, if the difference between the outlet pressure and control input port pressure is less than the pump margin, the pump will increase the displacement. If the difference between the outlet pressure and control input port pressure is greater than the pump margin, then pump displacement is reduced. It is commonly known that flow through an orifice can be represented as being proportional to the flow area and the square root of differential pressure. Since this pump control method provides a constant differential pressure of “pump margin”, the flow out of the primary pump 50 will be linearly proportional to the flow area between the pump outlet and load sense port 39. Alternatively, the primary pump 51 may be a type that has a positive displacement, non-positive displacement, electrohydraulically controlled displacement, or a load sense controlled displacement.

When multiple functions are demanding fluid the pump may be at a relatively high displacement that can overload the engine driving the pump and potentially cause the engine to stall. This condition is detected by the engine controller which responds by providing an alert signal to a system controller 57 for the hydraulic system, see FIG. 3. The system controller 57 responds by operating the load sense power control valve 37 which opens by proportional amount to reduce the pressure that is applied at in the load sense port 39 to control the outlet pressure of the primary pump 50. This action reduces the load on the engine and prevents stalling.

The system controller 57, in addition to receiving input signals from various sensors on the excavator, also receives signals from an operator interface 59 in the cab 11. The system controller responds by producing signals that operate the valves in the first hydraulic system 30.

Referring again to FIG. 2, each control valve 41-46 is an open-center, three-position valve, such as a spool type valve, for example. Although in the exemplary hydraulic system 30, the control valves 41-46 are indicated as being operated by a pilot pressure operated, one or more of them could be operated by a solenoid or a mechanical linkage.

The common features of all the control valves will be described with respect to the first control valve 41, then the features that are unique to the control valve in some of the hydraulic functions will be described. The first control valve 41 has a supply port 62 that is connected to the primary supply conduit 58 from the primary pump 50. A variable flow source orifice 64 within the control valve provides fluid communication between the supply port 62 and a flow outlet 66. The flow outlet 66 of the first control valve 41 is connected to a load sense conduit 67 by a function flow limiter 63 comprising a fixed orifice in parallel with a check valve. The second control valve has a similar function flow limiter 63, whereas the flow outlets 66 of the third through sixth control valves 44-46 are directly connected to the load sense conduit 67. A flow summation node 74 is defined in the load sense conduit 67. Thus, each variable flow source orifice 64 within a control valve provides a separate variable fluid path between the primary supply conduit 58 and the flow summation node 74.

The flow outlet 66 also is connected to a metering orifice inlet 70, either directly as for the first and second control valves 41 and 42 or by a conventional load check valve 68 as for the other control valves 43-46. A variable metering orifice 75 within the first control valve 41 selectively connects the metering orifice inlet 70 to one of two workports 76 and 78 depending upon the direction that the first control valve is moved from the center, neutral position. The two workports 76 and 78 connect to different ports on the associated hydraulic actuator, such as actuator 20 in the left travel function 31. The first control valve 41 is normally biased into the center position in which both workports 76 and 78 are closed.

The first control valve 41 also has a bypass orifice 80 that is directly connected between a bypass inlet 79 and a bypass outlet 81 of that control valve. The bypass orifices 80 for each of the other control valves 42-46 are connected in series to provide fluid communication between the summation node 74 and the return conduit 60. Specifically for the exemplary hydraulic system 30, the bypass inlet 79 of the fifth control valve 45 is connected to the summation node 74. The bypass outlet 81 of that control valve 45 is directly connected to the bypass inlet 79 of the fifth control valve 45 whose bypass outlet is directly connected to the bypass inlet 79 of the fourth control valve 44 and so on through all the hydraulic functions. The bypass outlet 81 of the first control valve 41 is connected directly to the return conduit 60. Thus the series of the bypass orifices 80 is connected between the summation node 74 and the return conduit 60.

Before describing the operation of the first hydraulic system with fluid from other pumps 82, 83 and 84, the control of the variable displacement primary pump 50 will be described. FIG. 3 illustrates those components of the first hydraulic system 30 that control the displacement of a primary pump 50. The variable flow source orifices 64 and the bypass orifices 80 are arranged in more functional groupings with those respective orifices shown outside the corresponding control valve 41-46 in which they are actually located. This functional diagram shows that the variable flow source orifices 64 ₄₁-64 ₄₆, of all the control valves 41-46 are connected in parallel between the primary supply conduit 58 from the primary pump 50 and the flow summation node 74. This parallel connection forms a variable flow section 65. The bypass orifices 80 ₄₁-80 ₄₆ of all the control valves 41-46 are connected in series between the flow summation node 74 and the return conduit 60 to the tank 51 and form a bypass section 88 of the hydraulic system 30. A subscript for an orifice's reference number denotes the control valve 41-46 of which the corresponding orifices is a part, e.g. bypass orifices 80 ₄₁ is part of the first control valve 41.

Assume initially that all the control valves 41-46 are in the center position in which both their workports 76 and 78 are closed off. In that state, the output from the primary pump 50, applied to primary supply conduit 58, passes through the variable flow source orifices 64 ₄₁-64 ₄₆, all of which are now shrunk to relatively small flow areas. Therefore, a relatively small amount of fluid flows from the primary pump 50 through the variable flow section 65 to the summation node 74. At this time, all the bypass orifices 80 ₄₁-80 ₄₆ in the bypass section 88 are enlarged to provide relatively large flow areas, thereby allowing the fluid entering the summation node 74 to pass easily into the return conduit 60. As a consequence, the pressure at the fluid summation node 74 is at a relatively low level. That low pressure level is transmitted through the load sense conduit 67 to the load sense port 39 of the variable displacement primary pump 50. This results in a low outlet pressure at the primary pump.

Alternatively when a control valve 41-46 is in the center position, its variable flow source orifice 64 ₄₁-64 ₄₆ can be fully closed so that no fluid flows through that control valve between the primary supply conduit 58 and the flow summation node 74. In this version of the pump control system, a separate small, fixed orifice may be added to connect the primary supply conduit 58 to the flow summation node 74 in the variable flow section 65, so that some flow from the primary supply conduit enters the flow summation node when all the control valves are in the center position.

Referring to FIGS. 2 and 3, operation of the pump control technique will be described in respect of the left travel function 31 with the understanding that the other hydraulic functions 32-36 operate in the same manner. The opening movement of the first control valve 41 in either direction from the center position connects the metering orifice inlet 70 through the variable metering orifice 75 to one of the workports 76 or 78, depending upon the direction of that motion. Opening the first control valve 41 also connects the other workport 78 or 76 to the return conduit 60. At the same time, the variable flow source orifice 64 ₄₁ enlarges by an amount related to the distance that the control valve moves, thereby causing the pump to increase fluid flow from the primary supply conduit 58 to the flow summation node 74 in order to maintain the “pump margin,” as previously described. Simultaneously, the size of the bypass orifice 80 ₄₁ shrinks, causing pressure at the summation node 74 to increase. Thus as the first control valve 41 opens a path through which fluid is supplied to the left travel hydraulic actuator 20, the flow through the variable flow section 65 into the summation node 74 increases, while the restriction, created by bypass orifice 80 ₄₁ to flow occurring out of that node to the tank return conduit 60 also increases thereby causing the pressure at the flow summation node 74 to increase.

When the flow summation node pressure is sufficiently great to overcome the load force acting on the left travel hydraulic actuator 20, fluid begins to flow through the metering orifice 75 in the first control valve 41 to drive the left travel hydraulic actuator 20.

At the same time that the first control valve 41 is opening one or more of the other control valves 42-46 also may be open. Their respective variable flow source orifices 64 ₄₂-64 ₄₆ also will be conveying fluid from the primary supply conduit 58 into the flow summation node 74. Because all the variable flow source orifices 64 ₄₁-64 ₄₆ are connected in parallel, the same pressure differential is across each of those orifices. That pressure differential and the cross sectional area of each flow source orifice determines the amount of flow through that orifice. The total flow into the flow summation node 74 is the aggregate of the individual flows through each variable flow source orifice 64 ₄₁-64 ₄₆. As a result, the sum of the areas that each variable flow source orifice is open determines the aggregate flow into the flow summation node 74 and thus controls the output flow from the variable displacement primary pump 50. The respective flow area of the metering orifice 75 in each control valve 41-46 and the respective load forces on actuators 17, 18, 19, 20, 22 and 26 determine the amount of flow each actuator receives from the flow summation node 74.

When the left travel hydraulic actuator 20 moves the excavator to the desired position, the first control valve 41 is returned to the center position by whatever apparatus controls that valve. In the center position, the two workports 76 and 78 are closed again cutting off fluid flow from the flow summation node 74 to the left travel hydraulic actuator 20. In addition, the associated variable flow source orifice 64 ₄₁ shrinks to a relatively small size which reduces the flow from the primary supply conduit 58 to the flow summation node 74. Returning the first control valve 41 to the center position also enlarges the size of its bypass orifice 80 ₄₁. Now if the other control valves 41-46 also are in the center position, all their bypass orifices 80 ₄₁-80 ₄₆ are relatively large, thereby relieving the flow summation node pressure into the return conduit 60.

Referring again to FIG. 2, a principal feature of the first hydraulic system 30 is the additional incorporation of one or more fixed displacement pumps 82, 83, and 84. Specifically, the left travel hydraulic function 31 has a first fixed displacement pump 82 that conveys fluid from the tank 51 into a first secondary supply conduit 91 which is connected to a secondary inlet port 92 on the first control valve 41. In the center or neutral, position of the first control valve 41, the secondary inlet port 92 is connected to a secondary outlet port 93 that in turn is coupled by a check valve 94 to a shared supply conduit 95.

The right travel function 32 includes a similar second fixed displacement pump 83 that furnishes pressurized hydraulic fluid to a second secondary supply conduit 96 connected to a second secondary inlet port 97 of the second control valve 42. That control valve has a second secondary outlet port 98 that is coupled by a check valve 99 to the shared supply conduit 95. Note that the two check valves 94 and 99 allow fluid to flow only from the respective secondary outlet port 93 or 98 into the shared supply conduit 95 and do not allow fluid to flow in the opposite direction.

The shared supply conduit 95 is connected directly to a third secondary inlet port 100 of the third control valve 43 for the swing function 33. As will be described, this connection provides a secondary source of hydraulic fluid to that third control valve 43 for use in operating the swing hydraulic actuator 26.

The boom function 34 includes a third fixed displacement pump 84 having an outlet that is connected to a tertiary supply conduit 102 that leads to a tertiary inlet port 104 of the fourth control valve 44. The fourth control valve 44 has a tertiary outlet port 106 that is directly connected to a secondary tank return conduit 110. In the center, or neutral, position of the fourth control valve the tertiary inlet port 104 is coupled directly to the tertiary outlet port 106 thereby allowing the outlet flow from the third fixed displacement pump 84 to flow directly to the secondary tank return conduit 110.

The displacement of the third fixed displacement pump 84 for the boom function 34 is selected to provide a predefined amount (e.g., 25 percent) of the total displacement of the pumps on a conventional hydraulic system for an excavator. For example, if a conventional excavator has two 100 cc pumps, then the third fixed displacement pump 84 would have a fixed displacement of 50 cc. Thus, when the fourth control valve 44 is moved into one of the open positions, fluid from the third fixed displacement pump 84 is always furnished to the boom hydraulic actuators 17.

The fixed displacement first and second fixed displacement pumps 82 and 83 have lower maximum output flow and pressure than the maximum flow and pressure available from the variable displacement primary pump 50 and are associated with hydraulic functions that normally require a significantly lower output pressure than that provided by the primary pump 50. For example, the variable displacement primary pump 50 may have a maximum output flow of 150 cc and the maximum output flow from each of the first and second fixed displacement pumps 82 and 83 may be 25 cc. The third fixed displacement pump 84 for the boom function may have a 50 cc maximum output flow, for example.

When the boom 13 is raising, all the fluid output from the third fixed displacement pump 84 is directed to the head chambers of the boom hydraulic actuators 17. Additional fluid from the primary supply conduit 58 also will be directed through the fourth control valve 44 to the head chambers of the boom hydraulic actuators 17. As noted previously, because of the heavy weight of the boom assembly 12, relatively high pressure is required to drive the boom hydraulic actuators 17 and raise the boom. In contrast, the arm and bucket functions 35 and 36 usually require significantly lower pressure even under their greatest load conditions. If other hydraulic functions consume all of the available output from the variable displacement primary pump 50, the boom 13 still will raise with the output from the third fixed displacement pump 84, albeit that motion is slower than if fluid also is available from the primary supply conduit 58. This use of a dedicated third fixed displacement pump 84 for the boom function 34 precludes the need to incorporate throttling losses in the lower pressure functions on the excavator.

On a conventional excavator, the travel functions are typically given priority for the available fluid produced from the shared pumps. This is because while travel motion is occurring, the movement of the boom, arm, and bucket are usually incidental to the travel. Thus on a previous excavator, the entire flow from one of the pumps was directed exclusively to the travel function, while the flow from another pump was utilized for the implement functions (boom, arm and bucket). Flow that was not consumed by the implement functions was routed back to the travel functions through a series of check valves and orifices. This type of system typically maintained one-half the machine's overall fluid displacement for the travel function regardless of the flow requirements of the implement functions. Although such a conventional circuit was effective, it did not provide priority to the boom up function, especially at low engine idle speeds, since the pressure allowed for the implement functions was limited by a fixed orifice. This previous hydraulic circuit also encountered controllability problems during a transition from a mode that combined the output from the two pumps to one in which one pump was used for travel and the other for the implement functions. The present circuit with separate first and second pumps for the travel function overcomes these disadvantages.

The first and second fixed displacement pumps 82 and 83 provide flow priority to each of the left and right travel functions 31 and 32, respectively, regardless of the flow consumption of the other hydraulic functions 33-36. With reference to FIG. 2, when the first control valve 41 is in one of the open positions, the flow from the first fixed displacement pump 82 is directed to the left travel hydraulic actuator 20. Any available flow from the primary supply conduit 58 also is directed through an internal orifice in the first control valve 41 to the left travel hydraulic actuator 20. That internal orifice and the orifice in the function flow limiter 63 restrict reverse flow of fluid produced by the first fixed displacement pump 82 through the first control valve and into the primary supply conduit 58, thereby robbing that fluid from being used in driving the left travel actuator 20.

Note also that when one or both travel functions 31 and 32 is not consuming fluid from its respective first or second fixed displacement pump 82 and 83, the associated control valve 41 or 42 conveys that fluid to the shared supply conduit 95 where it is available for powering the swing function 33. If the swing function is not operating, i.e. the third control valve 43 is in the center position, the fluid from the shared supply conduit 95 flows through that control valve into the tank return conduit 60. If, however, the swing function 33 is operating, the fluid available in the shared supply conduit 95 is applied along with fluid from the primary supply conduit 58 to drive the swing hydraulic actuator 26. This operation of the swing function is similar to how the travel functions 31 and 32 combine fluid from the first and second fixed displacement pumps 82 and 83 with fluid from the primary supply conduit 58 to drive the travel hydraulic actuators. When both travel functions 31 and 32 are being commanded 100%, the swing function 33 can only draw fluid from the primary pump 50 and as a consequence has its available flow limited. Because the swing function is in parallel with the travel functions, the pressure (and thus the torque) in the swing function also is limited.

Note that the load sense conduit 67 has a fixed orifice 69 between the connection points of that conduit for the two travel functions 31 and 32 and the conduit connection points for the remaining functions 33-36. With the first hydraulic system 30, the variable displacement primary pump 50 has a higher flow capacity than can be allowed into any one of the travel hydraulic actuators 20 or 22. As a consequence, limiting the application of the flow from the primary pump 50 to the travel hydraulic actuators is required. If the only function that is active is one of the two travel functions 31 or 32, there is no travel over speed issue since that travel function is in sole control of the primary pump and thus receives 75% of its flow requirements from the primary pump 50 and 25% of its flow from the first or second fixed displacement pump 82 or 83, in the exemplary system. It should be understood that other flow proportions can be employed in different systems.

However, when one of the travel functions 31 or 32 is active at the same time that at least one of the other functions 33-36 is active, that other function could demand a greater amount of flow from the primary pump 50, thereby increasing that pump's displacement to a point where the travel function could be driven into an undesirably high speed. To prevent this from occurring, the load sense conduit 67 has a limiting orifice 69 separating the conduit connections of the travel functions 31 and 32 from the conduit connections for the other hydraulic functions 33-36. This orifice 69 limits the degree to which those other functions can command the displacement of the primary pump 50, while allowing the active travel functions 31 and/or 32 to dominate that control. Thus, the additional flow that is allowed into the travel functions 31 and 32 is limited to flow created through the flow limit orifice 69 with the pump margin acting across it. For instance, if the flow limiting orifice 69 is sized to allow 25 lpm at 15 bar pressure (the margin), the maximum flow beyond the single function flow into the travel, is 25 lpm.

When both of the travel functions 31 and 32 are active, it is necessary to prevent more than the maximum travel flow from being directed to either travel hydraulic actuator 20 or 22. This is accomplished by the function flow limiter 63 connected between the flow outlet 66 of the control valve 41 and 42 and the load sense conduit 67 in those travel functions 31 and 32. This function flow limiter 63 comprises a fixed orifice in parallel with a check valve that forces fluid flowing in the direction from the flow outlet 66 to the load sense conduit 67 to pass through the fixed orifice. As an example, if both of the left and right travel functions 31 and 32 are commanded at 100% and then the left travel actuator 20 stalls and cannot consume its commanded flow, the flow will go through the orifice in the function flow limiter 63 for the left travel function 31 and into the load sense conduit 67. That flow continues through the load sense conduit 67 and the check valve in the function flow limiter 63 of the right travel function 32. As a result, a pressure differential appears across the orifice in the function flow limiter 63 of the left travel function 31 and thus the extra flow that is received by the right travel function 32 from the left function is limited by this orifice with a pressure differential equal to the margin pressure of the pump. Under typical conditions, this additional flow will be sufficiently small and does not result in over speed of the travel hydraulic actuators 20 and 22.

FIG. 4 illustrates a second hydraulic system 200 that embodies the present inventive concept. This hydraulic system 200 has a left travel function 201, and right travel function 202, a boom function 203, a swing function 204, an arm function 205, and a bucket function 206.

A variable displacement, primary pump 208 draws fluid from a tank 210 and furnishes that fluid under pressure into a primary supply conduit 209. The primary supply conduit 209 has a two-position proportional control valve 207 that couples a first section of that conduit, to which the left and right travel functions 201 and 202 are connected, to a second section of the primary supply conduit, to which the remaining hydraulic functions 203-206 are connected.

The second hydraulic system 200 has a fixed displacement pump 220 which also draws fluid from the tank 210 and furnishes that fluid under pressure through a supply check valve 222 to a boom/arm selector valve 224. The boom/arm selector valve 224 directs the output flow from the fixed displacement pump 220 into either a secondary supply conduit 228 or a bypass node 229. The bypass node 229 is connected by a check valve 231 to the load sense conduit 230. That check valve 231 prevents the flow from the fixed displacement pump 220 from flowing into the load sense conduit and thereby maintains the flow priority for the boom, swing, and arm functions in that priority order. Another check valve 233 allows fluid from the fixed displacement pump 220 that is not otherwise consumed by certain hydraulic functions to flow into the primary supply conduit 209 thus supplementing flow from the primary pump 208 for other hydraulic functions. This reduces the engine power drawn by the primary pump 208.

Each hydraulic function 201, 201, 203, 204, 205 and 206 respectively comprises a control valve 211, 212, 213, 214, 215 and 216 and the associated hydraulic actuator 20, 22, 17, 26, 18 and 19. All the control valves 211-216 are connected to the primary supply conduit 209 and to a return conduit 218 leading back to the tank 210. The control valves 211-216 are open-center, three-position types and may be a solenoid operated spool type valve, for example. Each the control valve 211-216 has two open states in which fluid from the primary supply conduit 209 is fed to the associated hydraulic actuator 17-26 and fluid from the actuator is returned through the valve to the tank return conduit 218. Depending upon which open state is used, the hydraulic actuator is driven in one of two directions.

The first and second control valves 211 and 212 have an supply port 221 that is directly connected to the primary supply conduit 209. An outlet port 223 of those control valves 211 and 212 is coupled by a function flow limiter 225 to a load sense conduit 230. The third, fifth and sixth control valves 213, 215 and 216 have similar supply ports 235 that are connected directly to the primary supply conduit 209 and outlet ports 236 that are connected directly to a load sense conduit section 238.

The fourth control valve 214 for the swing function 204 has its supply port 237 coupled by a proportional flow limit valve 246 to the primary supply conduit 209 and has an outlet port 239 that is connected directly to the load sense conduit section 238. Flow limit valve 246 is pilot operated by the pressure at the outlet port 239. The swing function 204 has a flow limiter that limits a magnitude of the flow from the variable displacement pump from exceeding the maximum flow rating for the swing hydraulic actuator 26. That flow limiter includes a flow valve 248 in series with a fixed orifice 250 through which fluid being supplied to the swing hydraulic actuator 26 travels. The flow valve 248 that is normally open and is pilot operated by the pressure differential across the orifice 250. Thus when the flow across the fixed orifice 250 exceeds a preset level, thereby producing a pressure drop of a given magnitude, the flow valve 248 begins to close proportionally thereby restricting the flow to the swing hydraulic actuator 26.

The load sense conduit section 238 is coupled by a fixed summation orifice 242 to the load sense conduit 230 in which a flow summation node 232 is defined. The load sense conduit 230 is coupled by a fixed orifice 241 to the displacement control input 234 of the primary pump 208. When a control valve 211-216 is open, fluid from the primary supply conduit 209 is applied to the flow summation node 232 and the amount of that fluid application is proportional to the degree to which the respective control is open.

The control valves 211-216 also have bypass orifices 240 that are connected in series to form a bypass passage 226 between the bypass node 229 and the tank return conduit 218. The bypass passage 226 along with check valve 231 also provide a fluid path between the summation node 232 and the return conduit 218. When all the control valves 211-216 are in the closed, center position, their bypass orifices 240 are enlarged to provide a relatively a large flow path which permits fluid to pass easily from the bypass node 229 to the return conduit 218. When a control valve 211-216 opens, its bypass orifice 240 shrinks restricting flow through the bypass passage 226 which causes pressure at the summation node 232 to increase, thereby increasing the fluid supply pressure.

Note that there are sets of dual check valves 255, 260 and 262 at control valves 213, 215, and 214, respectively. When the bypass passage 226 has a proper pressure therein, one of these check valves can open to supply fluid from the bypass passage to the respective control valve. The other check valve in the set prevents that fluid from flowing backwards into the load sense conduit 230 or into the primary supply conduit 209 in the open state of the respective valve. These check valves 255, 260 and 262 allow fluid from both the primary supply conduit 209 and the fixed displacement pump 220 to be supplied to a hydraulic function.

With continuing reference to FIG. 4, when either of the boom up or the arm in motions is commanded, the flow from the fixed displacement pump 220 is directed to the respective boom or arm function 203 or 205. This is accomplished by activating boom/arm selector valve 224 to proportionally direct the flow from the fixed displacement pump 220 into the secondary supply conduit 228. This prevents all of the fixed displacement pump flow from being consumed by the travel functions 201 or 202 and more importantly from being directed into the primary supply conduit 209 through the check valve 233. The flow in the secondary supply conduit 228 is directed into the bypass passage 226 through branch 253 at the boom function 203. Note that check valve 254 in the bypass passage 226 blocks this flow from traveling back to the bypass node 229. Thus, under all system conditions, if the boom function 203 is commanded, the flow from the fixed displacement pump 220 is directed with highest priority to maintain boom flow within the pressure limits of that function. In this case where a boom up operation is commanded, the bypass orifice 240 of the boom control valve 213 closes slightly, thereby forcing the fluid that has entered the bypass passage 226 to flow through check valve 255 and the boom control valve to the boom hydraulic actuators 17. This flow supplements any flow that would otherwise be drawn from the primary supply conduit 209.

Furthermore, during a digging operation of the excavator 10, when the arm function 205 is active, the boom/arm selector valve 224 also sends flow from the fixed displacement pump 220 into the secondary supply conduit 228. This flow also passes through the branch 253 into the bypass passage 226 and from there through to the arm function 205. Since the arm control valve 215 for that function has a reduced bypass orifice 240, the bypass passage flow is forced through a check valve 260 and the arm control valve to power the arm hydraulic actuator 18. It is quite common during a digging operation that the arm function 205 requires a higher pressure than the bucket function 206. The second hydraulic system 200 maintains the higher pressure from the fixed displacement pump 220 for the arm function, while the variable displacement primary pump 208 is allowed to run at a lower pressure as required by the bucket function 206.

Note that between the boom function 203 and the swing function 204, the bypass passage 226 is coupled through a check valve 256 and a fixed orifice 258 to the primary supply conduit 209. This circuit branch allows fluid that is not consumed by the arm function 205 to be directed into the primary supply conduit 209 from which it can be used by other hydraulic functions. Assuming that the boom function 203 and the swing function 204 are inoperative, when the arm function 205 is active, its bypass orifice 240 in control valve 215 is at least partially closed allowing fluid to flow into that function from the bypass passage 226 via the check valve 260. Any fluid that is not consumed by the arm function 205 flows through the check valve 256 and the fixed orifice 258. The fixed orifice 258 allows the pressure in the bypass passage 226 to be maintained so that the arm function will receive pressurized fluid.

When boom up, swing, and another lower pressure operation, such as arm in or bucket curl, are being commanded, the swing function 204 needs to maintain sufficient torque to accelerate properly. Under this command scenario, the output flow from fixed displacement pump 220 is directed to the boom function 203 via the secondary supply conduit 228 and that function thereby operates at the required pressure. The boom in or bucket curl operation are powered from the primary pump 208 at a lower pressure. The swing function 204, in order to accelerate, requires a higher pressure than the variable displacement primary pump 208 is producing. Therefore, the swing function 204 now is connected through the check valve and orifice combination 264 that directs some of the higher pressure flow in the secondary supply conduit 228 from the boom function 203 to the swing function. The size of orifice at 264 is selected to limit the flow that is diverted from the boom function.

Unlike the first hydraulic system 30, the second hydraulic system 200 does not have separate pumps for the left and right travel functions 201 and 202. Likewise, however, the variable displacement primary pump 208 has a significantly higher flow capacity than can be allowed into the travel hydraulic actuators 20 and 22 without an over speed condition occurring. When only one of the travel functions 201 and 202 is operating, it is in control of the primary pump 208 and thus receives the majority of its flow requirement from that pump. The remainder of the flow requirement is satisfied from the fixed displacement pump 220 via selector valve 224 and check valve 233 supplying that fluid into the primary supply conduit 209. When a single travel function is commanded along with an implement function, such as the bucket function 206, any additional flow to the travel functions 201 and 202 is limited by the fixed summation orifice 242 in the load sense conduit 230. As described previously with respect to the first hydraulic system 30, the same type of flow limiting occurs when both travel functions are active.

Because the travel functions 201 and 202 do not have separate fixed displacement pumps, a throttling priority technique is implemented in the second hydraulic system 200. In this instance, the control valve 207 separates the primary supply conduit 209 into a first section 270 to which only the travel functions 201 and 202 are connected and into a second section 272 to which the other functions 203-206 are connected. When a travel function is commanded, this control valve 207 transitions from an open position to a restricted position to limit the amount of flow allowed from the primary pump 208 to the non-travel functions 203-206. The control valve 207 closes proportionally to the highest pressure produced in the actuators for the two travel functions 201 and 202. In addition, the fixed summation orifice 242 in the load sense conduit 230 limits the amount of pump outlet flow commanded by the travel functions 201 and 202 that is allowed to flow to the implement functions 203, 205 and 206 during this mode of operation.

To avoid high pressure flow losses across the cross port relief valves 266 at the hydraulic actuator 26 of the swing function, a flow limit valve 246 is located in the flow path through the swing control valve 214 between the primary supply conduit 209 and the load sense conduit section 238. When the pressure in that load sense conduit section 238 rises above a preset level that is a little higher or a little lower than the cross port relief valve pressure threshold, the pilot operated control valve implementing this flow limit valve 246 closes to thereby limit the swing function's inlet flow from the primary pump 208. Note that the flow limit valve 246 may be placed on either the supply conduit side or the load sense conduit side of the swing control valve 214.

To improve productivity and match the pressure load of the bucket function 206 and the boom function 203, a throttling loss is added in the exhaust conduit of the bucket function between the control valve 216 and the tank return conduit 218. This restriction varies in proportion to the boom up command. In the second hydraulic system 200, this restriction is implemented by a proportional control valve 268 that is operated in response to the magnitude of the boom command. Alternatively, such a restriction could be implemented by a variable orifice on the boom spool through which the oil exhausting from the bucket function flows.

With reference to FIG. 5, a third hydraulic system 300 has six hydraulic functions, specifically a left travel function 301, a right travel function 302, a swing function 303, a boom function 304, an arm function 305, and a bucket function 306. The third hydraulic system 300 is similar to the first hydraulic system 30, except for the boom function 304. As a consequence, the common system components and elements have been assigned identical reference numerals in both FIGS. 2 and 5.

Instead of the dedicated pump 84, the boom function 304 incorporates a hydrostatic pump-motor 310 connected in parallel with the boom control valve 316. All the pumps 50, 82, 83 and 310 are driven by the engine of the excavator 10. The hydrostatic pump-motor 310 has a first port that is connected by a first check valve 311 to the tank return conduit 60, wherein the first check valve allows fluid to flow only in a direction from the tank return conduit 60 to that pump port and to the rod chamber of the boom hydraulic actuators 17. The first port of the hydrostatic pump-motor 310 also is connected directly to a first workport 314 of the boom control valve 316. A second port of the hydrostatic pump-motor 310 is directly connected via an actuator conduit 319 to a second workport 315 of the boom control valve 316. A second check valve 312 is connected to allow fluid to flow from the tank return conduit 60 to flow to the second port of the hydrostatic pump-motor 310 and to the second workport 315. The first and second check valves 311 and 312 provide an anti-cavitation function ensuring that sufficient make-up fluid is provided to maintain the hydrostatic pump-motor 310 filled with fluid. Those check valves 311 and 312 can be removed if there otherwise is sufficient make-up fluid to maintain that filled condition. The first workport 314 is connected to the rod chambers of the boom hydraulic actuators 17 and the second workport 315 is coupled by a load check valve 318 to the head chambers of those hydraulic actuators. The load check valve 318 is pilot operated into an open state by a pressure signal when the boom control valve 316 is in the position in which fluid is to drain from those head chambers.

The control valve 316 for the boom function 304 is operated by pilot pressures applied to opposite ends of that control valve and those pilot pressures also are applied to a port of a displacement control device 317 for the hydrostatic pump-motor 310. Alternatively, electrical actuators can be used to operate both the control valve 316 and the displacement control of the hydrostatic pump-motor 310.

During a boom raise operation, the pilot signal that operates the control valve 316 also controls the displacement of the hydrostatic pump-motor 310. The initial supply fluid flow to the boom hydraulic function 304 is provided via the control valve 316 from the flow summation node 74 in the same manner as that described with respect to the first hydraulic system 30. As the boom function 304 continues to be commanded, the hydrostatic pump-motor 310 functions in the pumping mode in which fluid exiting the rod chambers of the hydraulic actuators 17 is forced into the head chambers. In the pumping mode, the fluid exiting the ports for the head chambers of the hydraulic actuators is at a lower pressure than is required for fluid to enter the port for the rod actuator chamber. Therefore the hydrostatic pump-motor 310 must function as a pump to increase the pressure of the fluid flowing through that device.

If one or more other hydraulic functions 301-303 or 305-306 is operating simultaneously at a lower pressure and consuming fluid at a higher flow rate, there may be inadequate fluid available in the primary supply conduit 58 for the boom hydraulic function 304. When this condition occurs, the hydrostatic pump-motor 310 may be the only source of fluid for the boom hydraulic function. Nevertheless, priority for the boom operation is maintained with this hydraulic circuitry.

When a boom lower operation is commanded, the same pilot signal that operates the boom control valve 316 also controls the displacement of the hydrostatic pump-motor 310. In this mode, the initial exhaust flow from the head chambers of the boom hydraulic actuators 17 is governed by the boom control valve 316, which directs that flow directly into the tank return conduit 60, with no recovery of the energy in that fluid. Note that at this time, the load check valve 318 is forced open by the pilot pressure signal operating the boom control valve 316. As the machine operator commands greater boom motion and the pilot pressure increases, the hydrostatic pump-motor 310 begins operating in the motoring mode and consumes some of the fluid exhausting into the actuator conduit 319 from the head chambers of the boom actuators 17. The fluid that flows through the hydrostatic pump-motor 310 is sent into the expanding rod chambers of those actuators. This action causes the hydrostatic pump-motor 310 to begin motoring, thereby applying mechanical energy onto the drive shaft connected to the hydrostatic pump-motor and recovering energy from the exhausting fluid. The rate of fluid consumption by the hydrostatic pump-motor 310 is related to the magnitude of the operator commands and the resulting pilot pressure signal. Since the head chamber exhaust flow from the hydraulic actuators 17 is greater than the flow that can be consumed by the hydrostatic pump-motor 310, the remainder of that fluid flow is diverted to the tank 51 through the boom control valve 316.

During boom lowering when the bucket 15 is driven downward into the ground and the boom assembly 12 becomes a powered load, the hydrostatic pump-motor 310 no longer recovers energy. Instead the hydrostatic pump-motor 310 delivers energy to its port that is connected to the rod chambers of the boom hydraulic actuators 17 without a change in the pump-motor displacement. Thus a smooth transition occurs from motoring to pumping.

The foregoing description was primarily directed to one or more embodiments of the invention. Although some attention has been given to various alternatives within the scope of the invention, it is anticipated that one skilled in the art will likely realize additional alternatives that are now apparent from disclosure of embodiments of the invention. Accordingly, the scope of the invention should be determined from the following claims and not limited by the above disclosure. 

1. A hydraulic system for a machine comprising: a plurality of hydraulic actuators, including a given hydraulic actuator that has a cylinder and piston arrangement with a first port and a second port; a first pump for furnishing fluid for operating the plurality of hydraulic actuators; a tank for receiving fluid from the plurality of hydraulic actuators; a plurality of control valves, each of which selectively controls fluid flow between an associated one of the plurality of hydraulic actuators and both the first pump and the tank, the plurality of control valves including a given control valve having a first workport connected to the first port of the given hydraulic actuator and a second workport connected to the second port; and a hydrostatic pump-motor operatively connected to convey fluid between the first and second ports of the given hydraulic actuator, and having a motoring mode in which the hydrostatic pump-motor is driven by fluid flowing out of one of the first and second ports at a higher pressure than fluid flowing into the other one of the first and second ports, and having a pumping mode in which fluid is pumped by the hydrostatic pump-motor when fluid flowing out of one of the first and second ports has a lower pressure than is required for fluid to flow into the other one of the first and second ports.
 2. The hydraulic system as recited in claim 1 wherein the first pump has a variable displacement.
 3. The hydraulic system as recited in claim 1 wherein the hydrostatic pump-motor has a variable displacement.
 4. The hydraulic system as recited in claim 3 wherein the given control valve is operated by a control signal; and the variable displacement of the hydrostatic pump-motor is controlled by the control signal.
 5. The hydraulic system as recited in claim 1 wherein when the hydrostatic pump-motor is in the motoring mode, the given control valve conveys some of the fluid flowing out of the one of the first and second ports to the tank.
 6. The hydraulic system as recited in claim 1 wherein when the hydrostatic pump-motor is in a pumping mode, the given control valve conveys fluid from the first pump into the other one of the first and second ports.
 7. The hydraulic system as recited in claim 1 further comprising: a flow summation node in fluid communication with a control port for controlling displacement of the first pump; and wherein each of the plurality of control valves comprises a metering orifice for varying fluid flow between the flow summation node and the associated hydraulic actuator, a variable flow source orifice, and a variable bypass orifice, the variable flow source orifices of the plurality of control valves being connected in parallel between the first pump and the flow summation node, and the variable bypass orifices of plurality of control valves being connected in series to form a bypass passage through which fluid flows between the flow summation node and the tank.
 8. The hydraulic system as recited in claim 7 wherein in each of the plurality of valves, the variable flow source orifice enlarges as the metering orifice enlarges, and the variable flow source orifice shrinks as the metering orifice shrinks.
 9. The hydraulic system as recited in claim 7 wherein in each of the plurality of valves, the variable bypass orifice shrinks as the metering orifice enlarges, and the variable bypass orifice enlarges as the metering orifice shrinks.
 10. A hydraulic system for a machine comprising: a plurality of hydraulic actuators, including a given hydraulic actuator that has a cylinder and piston arrangement with a first port and a second port; a first pump for furnishing fluid for operating the plurality of hydraulic actuators; a tank for receiving fluid from the plurality of hydraulic actuators; a plurality of control valves, each of which selectively controls fluid flow between one of the plurality of hydraulic actuators and both the first pump and the tank, the plurality of control valves including a given control valve having a first workport connected to the first port of the given hydraulic actuator and a second workport connected to the second port; and a hydrostatic pump-motor operatively connected to convey fluid between the first and second ports of the given hydraulic actuator, and having a motoring mode in which the hydrostatic pump-motor is driven by fluid flowing out of one of the first and second ports at a higher pressure than fluid flowing into the other one of the first and second ports.
 11. The hydraulic system as recited in claim 10 wherein when the hydrostatic pump-motor is in the motoring mode, the given control valve conveys some of the fluid flowing out of the one of the first and second ports to tank.
 12. The hydraulic system as recited in claim 10 wherein the hydrostatic pump-motor further has a pumping mode in which fluid is pumped by the hydrostatic pump-motor when fluid flowing out of one of the first and second ports has a lower pressure than is required for fluid to flow into the other one of the first and second ports.
 13. The hydraulic system as recited in claim 12 wherein when the hydrostatic pump-motor is in a pumping mode, the given control valve conveys fluid from the first pump into the other one of the first and second ports.
 14. The hydraulic system as recited in claim 10 wherein the hydrostatic pump-motor is a variable displacement type.
 15. The hydraulic system as recited in claim 14 wherein the given control valve is operated by a control signal; and displacement of the hydrostatic pump-motor is controlled by the control signal.
 16. The hydraulic system as recited in claim 10 further comprising a pilot operated load check valve that opens in response to a pilot signal to allow fluid flow from the given hydraulic actuator to the second workport and otherwise allows fluid flow only from the second workport to the given hydraulic actuator.
 17. The hydraulic system as recited in claim 10 further comprising: a flow summation node in fluid communication with a control port for controlling displacement of the first pump; and wherein each of the plurality of control valves comprises a metering orifice, a variable flow source orifice, and a variable bypass orifice, wherein the metering orifice varies fluid flow between the flow summation node and the associated hydraulic actuator, the variable flow source orifices of the plurality of control valves being connected in parallel between the first pump and the flow summation node, and the variable bypass orifices of plurality of control valves being connected in series to form a bypass passage through which fluid flows between the flow summation node and the tank.
 18. The hydraulic system as recited in claim 17 wherein in each of the plurality of valves, the variable flow source orifice enlarges as the metering orifice enlarges, and the variable flow source orifice shrinks as the metering orifice shrinks.
 19. The hydraulic system as recited in claim 17 wherein in each of the plurality of valves, the variable bypass orifice shrinks as the metering orifice enlarges, and the variable bypass orifice enlarges as the metering orifice shrinks.
 20. A hydraulic system for a machine comprising: a plurality of hydraulic actuators, including a given hydraulic actuator that has a cylinder and piston arrangement with a first port and a second port; a first pump for furnishing fluid for operating the plurality of hydraulic actuators; a tank for receiving fluid from the plurality of hydraulic actuators; a plurality of control valves, each of which selectively controls fluid flow between one of the plurality of hydraulic actuators and both the first pump and the tank, the plurality of control valves including a given control valve having a first workport connected to the first port of the given hydraulic actuator and a second workport connected to the second port; and a hydrostatic pump-motor operatively connected to convey fluid between the first and second ports of the given hydraulic actuator, and having a pumping mode which pumps fluid when fluid flowing out of one of the first and second ports has a lower pressure than is required for fluid to flow into the other one of the first and second ports.
 21. The hydraulic system as recited in claim 20 wherein when the hydrostatic pump-motor is in the pumping mode, the given control valve conveys fluid from the first pump into the other one of the first and second ports.
 22. The hydraulic system as recited in claim 20 wherein the hydrostatic pump-motor further has motoring mode in which the hydrostatic pump-motor is driven by fluid flowing out of one of the first and second ports at a higher pressure than fluid flowing into the other one of the first and second ports.
 23. The hydraulic system as recited in claim 22 wherein when the hydrostatic pump-motor is in the motoring mode, the given control valve conveys some of the fluid flowing out of the one of the first and second ports to tank.
 24. The hydraulic system as recited in claim 20 wherein the hydrostatic pump-motor is a variable displacement type.
 25. The hydraulic system as recited in claim 24 wherein the given control valve is operated by a control signal; and displacement of the hydrostatic pump-motor is controlled by the control signal.
 26. The hydraulic system as recited in claim 20 further comprising a pilot operated load check valve that opens in response to a pilot signal to allow fluid flow from the given hydraulic actuator to the second workport and otherwise allows fluid flow only from the second workport to the given hydraulic actuator.
 27. The hydraulic system as recited in claim 20 further comprising: a flow summation node in fluid communication with a control port for controlling displacement of the first pump; and wherein each of the plurality of control valves comprises a metering orifice for varying fluid flow between the flow summation node and the associated hydraulic actuator, a variable flow source orifice, and a variable bypass orifice, the variable flow source orifices of the plurality of control valves being connected in parallel between the first pump and the flow summation node, and the variable bypass orifices of plurality of control valves being connected in series to form a bypass passage through which fluid flows between the flow summation node and the tank.
 28. The hydraulic system as recited in claim 27 wherein in each of the plurality of valves, the variable flow source orifice enlarges as the metering orifice enlarges, and the variable flow source orifice shrinks as the metering orifice shrinks.
 29. The hydraulic system as recited in claim 27 wherein in each of the plurality of valves, the variable bypass orifice shrinks as the metering orifice enlarges, and the variable bypass orifice enlarges as the metering orifice shrinks. 